Liquid pump



March 12,1963

Filed Feb. 8. 1960 J. C. FISHER LIQUID PUMP '7 Sheets-Sheet 1 Fig. l

INVENTOR. JOHN C; FISHER 'uzmww. JENNEY, WITTER & HILDRETH ATTORNEYS March 12, 1963 J. c. FISHEFQ 3,080,822

LIQUID PUMP Filed Feb. 8. 1960 '7 Sheets-Sheet 2 N '2" LL.

- JOHN c. FISHER KENWAY, 'JENNEY, WITTER & HILDRETH ATTORNEYS March 12, 1963 Filed Feb. 8. 1960 Fig. 3

J. C. FISHER LIQUID PUMP 7 Sheets-Sheet 3 mmvmx. JOHN c. FISHER BY KENWAY. 'JENNEY, WITTER &' HILDRE'H ATTORNEYS March 12, 1963 J. c. FISHER LIQUID PUMP 7 Sheets-Sheet 4 Filed Feb. 8. 1960 I Fig. 4

INVENTOR. JOHN C. FISHER KENWAY, JENNEY, WlTl'ER & HILDRETH ATTORNEYS J. c. FISHER LIQUID PUMP March 12, 1963 7 Sheets-Sheet 5 Filed Feb. 8. 1960 -MMMMMMI lvvlllflllflllllvy mmvrox. JOHN c. FISHER KENWAY. JENNEY. WITTER & HILDRETH ATTORNEYS March 12, 1963 Fig. 7

J. c. Flsl-lzia 3,080,822

LIQUID PUMP 7 Sheets-Sheet 6 Fig. 8 Fig. 9

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5s 72 72 52 -46 83 83 82 k -s2 82 M g INVENTOR. JOHN C. FISHER BY KENWAY, IENNEY, WITTER &' HILDRETH ATTORNEYS March 12, 1963 JQC. FISHER 3,080,822

LIQUID PUMP Filed Feb. 8. 1960 '7 Sheets-Sheet '7 Fig. l0

INVENTOR. JOHN C. FISHER KENWAY, 'JENNEY, WETTER & HILDRETH ATTORNEYS Unite States Patent'O acsaszz LIQUID PUMI John C. Fisher, Cambridge, Mass, assignor to Am Dyue Trust, a trust of Massachusetts Filed Feb. 8, 1960, Ser. No. 9,156 12 Claims. (Cl. 103-53) This invention relates to liquid pumps of the doubleacting piston type and particularly to an improved type which is capable of pumping corrosive, toxic, explosive, or otherwise dangerous liquids, without the leakage associated with conventional piston pumps which operate at low frequencies of reciprocation. Such pumps require comparatively long strokes in order to provide the requisite discharge; they'require piston rings or packings in order to minimize leakage around the piston; and they must have sliding-contact shaft seals where the shaft emerges from the body of the pump. These features make the ordinary piston pump unsuitable for use with corrosive liquids or highly active solvents, which, respectively, ruin precise ring-to-cylinder fits by erosion of the carefully machined mating surfaces, and leak through even the most ingenious seals and pack-ings now available. Because of the long stroke, a pump of this sort cannot employ a shaft seal of the continuous-enclosure type, and because of the low frequency of operation, it is not feasible to employ a piston without rings or packing because the volumetric efliciency would be low.

This application is a continuation-in-part of my prior copending application Serial No. 707,531, filed January 7, 1958, now abandoned.

The principal object of my present invention is to provide a double-acting piston pump which can pump corrosive, toxic, explosive, or otherwise dangerous liquids Without external leakage and with high volumetric efficiency, which requires no piston rings or packings, and which permits the discharge to be varied without the use of external control valves.

Further objects relate to certain structural and functional features which will be better understood from a consideration of the following disclosure.

In accordance with the present invention I provide a pump which comprises a housing in which is slidably mounted a piston with at least substantially all its side ,walls or wall being smooth and being in closely spaced out-cf-contact relation to the adjacent smooth side ,wall or walls of the housing. The piston may be supported by spaced longitudinally extending ribs projecting from its side wall or walls so as to provide the desired clearance, which need merely be such as to prevent contact between the adjacent surfaces. If desired, resilient 3,980,822 Patented Mar. 12, 1963 The piston is oscillated at a frequency between 20 and 300 cycles per second and to this end any suitable means may be employed, depending on the particular supporting structure and the size or capacity of the pump. For example, where the piston is slidably supported on or by means entirely within the confines of the housing, a

. magnetornotive member such as a coil surrounding the supporting fingers may be secured to the ends of the piston, or it may be slidable on a fixed rod within the housing, or it may be rigidly secured to a rod slidably mounted at or adjacent to the ends of the housing. In any event, the piston divides the housing into two chambers each having an inlet port and an outlet port.

7 The inlet ports are connected with an inlet manifold or header and are provided with suitable rectifying valve means operative to permit a liquid flow only from the in let manifold into the respective chambers. Similarly the outlet ports are connected with an outlet manifold or header and are provided with rectifying valve means operative to permit a liquid flow only from the respective chambers into the outlet manifold. An outlet or discharge duct is connected to the outlet manifold and an inlet duct or intake is connected with the inlet manifold, the discharge and intake being connected with an exterior circuit as in my copending application Serial No. 553,015, filed December 14, 1955, now Patent No.

housing maybe employed, but if the piston is rigidly supported on a piston rod extending beyond the confines of the housing then either an electrodynamic vibrator or other reciprocating motor may be employed.

The preferred construction of my pump comprises a cylindrical piston rigidly mounted on a straight, coaxial rod, the piston being located concentrically in a cylinder of the same shape but of slightly larger cross-sectional area, so that there is no point at which the piston touches the bore of the cylinder. The transverse crosssection of the piston and cylinder may be circular, triangular, square, rectangular, or any regular polygon. The cylinder is formed within a block and communicates at each end with a somewhat enlarged chamber. In order to avoid the rapid corrosion which occurs between piston rings and cylinder wall'when a conventional piston pump is used with corrosive liquids, the piston must be mounted so that it cannot touch the cylinder bore at any point, as has been mentioned above. However, this construction has the disadvantage that it creates a leakage space between the piston and cylinder walls, through which liquid will flowfrom the chamber which is momentarily at high pressure to the chamber which is momentarily at low pressure. This internal leakage flow is undesirable, because it reduces the external discharge from the pump. Several methods have been employed in the past to reduce this leakage flow to tolerable levels. Oneprior technique employed by L. S. Barengueras in United States Patent No. 1,233,438 comprises the use of perforations, or undulous grooves and ridges, in the surfaces of the piston and cylinder wall. The high cost of manufacturing such surfaces limits this method to very specialized applications Where cost is of little importance and has undoubtedly contributed to the lack of commercial use of this method. The essential feature of the Barengueras device is that it causes added turbulence and energy losses in the liquid flowing through the clearance space. These energy losses are functions of the velocity of the liquid in the clearance space as well as of the viscosity of the liquid. The Barengueras method becomes less effective as the viscosity of the liquid decreases and would be completely ineffectual with a fluid having zero viscosity. No such fluid exists, of course, but most gases approximate this condition.

Another technique is that employed by E. T. Booth, lit, in United States Patent No. 2,668,656, namely the use of a piston and cylinder wall which have cooperating surfaces forming a convoluted leakage path around the piston so that the leakage flow through the clearance space must undergo one or more rather abrupt or severe changes in direction. The surfaces of piston and cylinder wall in the vicinity of these bends in the passage are essentially smooth, being cylindrical or conical in their geometry. Such surfaces, while being cheaper to manufacture than those used by Barengueras, and while being reasonably effective in limiting the leakage, also impose an undesirable restriction on the stroke which the piston can undergo without striking any portion of the cylinder; The essential feature of the method of the Booth patent is the same as that of the Barengueras patent, namely the increase of energy losses in the leakage stream because of velocity-dependent eflects. It is, of course, a well known principle of fluid mechanics that loss of head from a flowing fluid stream because of friction is dependent upon the geometry of the flow passageway, the kinematic viscosity of the liquid, and the velocity of the liquid.

The preferred construction of my pump utilizes only the simplest and most economical geometry for the piston and cylinder bore, namely a straight, circular cylinder with smooth, unconvoluted surfaces spaced from a'smooth walled piston of smaller diameter. Since the annular unconvoluted leakage space so created is inherently low in its resistance to leakage flow, something else must be done to minimize leakage. In order to understand how this leakage flow can be controlled, we must examine the basic differential equation of the flow, under certain simplifying assumptions, which are fairly well fulfilled in practice. The assumptions are as follows:

(1) The motion of the piston is a regular, periodic function of time (sinusoidal, for example).

(2) The cylinder block is stationary.

(3) The liquid is incompressible.

(4) The variation of absolute pressure in each chamber at the ends of the piston is a regular, periodic function of time.

(5) The piston and bore are coaxial, and their common axis is horizontal, or nearly so. (It does not matter if the piston be vertical, as long as the maximum pressure head in each chamber is large by comparison with the length of the clearance space in the axial direction.)

(6) The change in elevation from top to bottom of the piston (assumed horizontal) is negligible by comparison with the maximum pressure head in each pump chamber.

(7) The clearance space is of constant cross-section in planes normal to the common axis, its bounding surfaces are smooth, and its radial dimension is small by comparison with its axial length.

(8) The liquid is Newtonian (i.e., the shear stress within the liquid is proportional to the rate of shear), and its temperature is constant.

Under the influence of the instantaneous pressure difference between the chambers at the ends of the piston, liquid will flow through the clearance space from the high-pressure chamber to the low-pressure chamber. Because the geometry of the clearance annulus is constant, and because the pressure difference and piston velocity are periodic functions of time, the leakage flow will itself be a regular, periodic function of time. It will have a zero average value over any whole number of cycles of piston motion. Therefore, the leakage flow is a purely alternating motion of liquid through the clearance space, and it will have both an alternating velocity and an alternating acceleration. It will become apparent from the subsequent discussion that this acceleration is the factor which is used in my pump to limit the leakage flow to a tolerable level.

During any half cycle of the leakage flow, the pressure gradient in the axial direction through the leakage annulus must have two components: (1) a component due to the acceleration of the liquid, and (2) a component due to the viscous shearing of the liquid. Under our foregoing assumptions, the pressure gradient at any instant must be constant over the whole axial span of the clearance space, and each plane normal to the piston axis must be a surface of constant pressure at any instant. Since both frictional and inertial forces act on the liquid, the velocity profile across the radial dimension of the leakage space will not be flat, but rather the velocity of the liquid at any instant will be zero at the cylinder wall, rise to some maximum value in a direction opposite to the piston velocity near the center of the annulus, and fall through zero velocity down to the value of piston velocity at the piston surface. It must be remembered that the piston forms one boundary of the leakage space, and this is a boundary which, in general, moves with a velocity opposite to that of the bulk of the leakage flow.

This velocity profile will change from instant to instant,

where =the shear stress along any cylindrical surface within the liquid and coaxial with the piston;

',u=the absolute viscosity of the liquid,

V =velocity of the liquid in the axial direction;

n=distance measured normal to the piston axis (radially).

It may be shown that the resulting pressure gradient is given by as or? on where P=the absolute pressure in the liquid; S=distance measured parallel to the piston axis.

where w=the specific weight of the liquid; g=the acceleration due to the earths gravity; t=time, measured from any convenient reference instant.

Note that the pressure gradient given by Equation 3 is in such a direction as to oppose the acceleration of the liquid (i.e. the pressure falls in the direction of positive acceleration, so that a pressure difference or drop is required to accelerate a liquid column over any finite distance parallel to the acceleration).

When both frictional (viscous) and inertial forces act simultaneously, as they do in this case, the net pressure gradign; must be the sum of those given by Equations eso75 6t (4) This is the basic differential equation of the leakage flow through the annulus between a cylindrical piston with smooth wall and a coaxial cylindrical bore with smooth wall. In general, it is very difiicult to solve this equation, but a qualitative study of the equation, and laboratory experiments on the physical situation, reveal certain pertinent facts:

(1) Since the piston velocity and the pressure difference are regular, periodic functions of time, the leakage fiow must be a regular, periodic function of time, describable by a Fourier series of time-harmonic functions having a fundamental (base) frequency equal to that of the piston velocity.

(2) If other parameters are held constant, the crest value and root-means-square value of the alternating leakage fiow will diminish as the fundamental frequency of the piston velocity is increased. Therefore, the total volume of liquid which passes through the leakage annulus on each discharge stroke of the piston, and thus diminishes both the outflow from the discharge valve on the fact that, for a given maximum liquid velocity, the maximum liquid acceleration increases with frequency, while the available pressure difference which causes the leakage does not increase correspondingly.

(3) In view of item (2) above, it is evident that the leakage flow through the annular clearance space can be decerased to any desired level merely by making the piston frequency high enough, and this method is effective even for liquids having negligible viscosity, provided only that they have a specific weight greater than zero. The heavier the liquid, the more effective is this method of limiting the leakage flow, but it is effective. even for the lightest liquids which are pumped in industry.

Experiment-s conducted upon actual pumps of the type described herein have shown that, at a piston frequency 60 cycles per second, with a piston 5 inches long, 3 inches in diameter, and having a radial clearance in. the bore of 0.010 inch, the equivalent lost external discharge due to the leakage how is 2.90 gallons of water per minute at an equivalent average pressure difference of 40 pounds per square inch, gage. The equivalence between average leakage flows and pressure differences, and their corresponding peak values was based upon the sinusoidal variations of these parameters which were used in conducting the tests:

Equivalent average value=:2 (peak value of parameter) By comparison with this leakage flow through the clearance space under dynamic conditions at 60 cycles per second, the steady leakage flow through the same annular space under constant pressure difference (i.e. zero fre quency) is 2.90 gallons per minute at only 3.50 pounds per square inch, gage. Conversely, at 40 p.s.i.g. the steady leakage would be 16.2 gallons per minute.

Furthermore, experiments of this same type show that i when the piston frequency is below 20 cycles, the leakage flow past the piston becomes undesirably large, even if the radial dimension of the clearance space is reduced to the point where it is difficult to maintain the out-ofcontact relation between piston and bore. On the other hand, if the piston frequency is made very high, the axial length of the clearance space becomes an appreciable fraction of the wavelength of sound in the enclosed liquid, and resonance phenomena can occur which actually magnify the leakage fio-w. Therefore, the preferred range of operation of my pump is 20 to 300 cycles per second. This operating frequency permits the use of the simplest possible geometry for both piston and bore: coaxial cylintiers whose cross sections are regular polygons of the same number of sides (preferably the circle, which is a regular polygon with an infinite number of sides) and whose surfaces are smooth and unconvoluted.

Each of the pump chambers at the ends of the piston is preferably provided wtih a pair of reed-type fluid rectifiers of the type described in my copending application Serial No. 636,597, filed January 28, 1957, now abandoned; one valve permitting liquid to enter the chamber but not to leave it, and the other valve permitting liquid to leave but not to enter. Within each. chamber, directly opposite the respective end of the cylinder bore, there is provided an opening through which the rod passes. This rod opening has the same cross-sectional shape (circular being the preferred shape) and a slightly larger cross-sectional area than the rod 'so that the shaft does not touch the bore of the shaft-hole at any point. 1

At each end of the structure, where the rod emerges :from the housing, there is provided a solid, but longir of suitable flexible tubing or hose.

tern-a1 to the rod is a space into which the liquid within the pump may flow through the small annular clearance space around the rod in the rod hole, but from which it cannot escape. to the outside environment.

A further extension of the rod at each end beyond the sealing boot or bellows providesv means for attachment ofthe rod to special springs or fiexures, which serve to. restrain the. rod against all motions other than those parallel to its axis. These flexures support the weight of the rod. and piston, thereby preventing it and the piston from touching their respective bores during operation so that no rubbing of solid surfaces against each other takes place.

Each of the two discharge valves communicates on its downstream side with the interior space of a discharge header or manifold which is attached to the cylinder block. Emerging from this discharge header is an outlet duct or nipple to which may be clamped a suitable flexible hose or tube, this tube serving to carry the emerging liquid flow to an external circuit or apparatus. Such a tube is preferably elastic so as to eliminate for the most part the pulsations inherent in the output, stream. The tube performs this latter function by virtue. of its elasticity, i.e., its ability toexpand and contract slightly in the radial direction in response to changes of internal pressure. For the frequencies of operation and amplitudes encountered in conventional piston pumps, this means of pulsation removal would not be so effective; but for higher frequencies, the distributed compliance of the tube will permit substantial velocity differences for the flow along its length with only a moderate fluctuation of internal pressure. Thus, as the frequency of operation increases, a flexible hose or tube becomes a progressively better pulsation eliminator, up to the point at which the length of the tube becomes an appreciable fraction of the wave length of sound in the enclosed liquid. The preferred frequency of operation of my pump is 60 cycles per second, the practical range of operating frequency being from 20 to 300 cycles per second.

Each of the two intake valves communicates on its upstream side with the interior space of an intake manifold or header which is attached to the cylinder block. Connected to the intake header is an intake duct or nipple which serves as a connection point for another length This tubing carries the entering flow from the external apparatus and may also serve to introduce into the steady flow from this apparatus the pulsations which were eliminated by the hose on the discharge side. Were these pulsations not introduced, cavitation would occur in each pump chamber on the intake stroke. However, the intake tubing prevents cavitation by means of slight radial expansions and contractions as the internal pressure fluctuations cyclically below the atmospheric pressure which surrounds the outer surface of the hose. If the operating dischage pressure is too high for the hose to resist wtihout reinforcement, it may be made with a woven fabric molded into the elastomer of the wall, or it may be made with the wall of pure elastomer and covered externally with a close-fitting woven metal braid.

One end of the shaft of my pump isconneoted through a drive link to the armature of an electrodynarnic vibration motor, which produces a reciprocating motion, the double amplitude of which does not exceed 0.75 inch.

This motion is customarily sinusoidal in its time variation,

whose. connecting rod actuates a crosshead attached to the pump shaft. However, the function of the pump is not dependent upon the specific details of the drive system.

In operation, the pump causes the liquid to flow because on each stroke of the piston the volume of the chamber at the trailing end is increased while that of the chamber at the leading end is decreased; on the succeeding stroke these events are interchanged for the two chambers. Hence, in each complete cycle, each chamber alternately receives an influx of liquid from the intake header and then discharges this excess liquid to the discharge header. The action of the valves prevents any significant backflow through the valve ports. Because the clearance space around the piston is small in area compared with the cross-sectional area of the piston itself, and because the period of each stroke is short, the leakage flow around the piston is effectively restricted. Both frictional forces due to viscous shearing of the liquid and inertial forces due to high acceleration of the liquid act to limit the leakage flow to a small fraction of the external discharge. Thus, the pump does not need the piston rings of conventional piston pumps in order to atain high volumetric efficiency. The leakage flow is further reduced by making the axial length of the piston as long as is feasible.

Although the space within the seal boot or bellows at each side of the cylinder block is open to liquid flow to and from the corresponding pump chamber, this leakage flow (leakage only in the sense of reducing the volumetric efliciency) is effectively limited by the following factors: (1) the small clearance space around the shaft, (2) the fact that the change in internal volume of the space within the boot is much less than the change in volume of the pump chamber, and (3) the fact that the intra-boot space may be reduced by placing a set of close-fitting elastomer washers on the portion of the shaft inside the boot.

In practice the piston, shaft, cylinder block, valve blocks, and headers are preferably made of a suitable metal, such as stainless steel. The intake and discharge ducts, together with the seals, may be made of a variety of elastomers such as polyvinylchlorides (Tygon), chlorinated polyethylenes (Hypalon), neoprenes, rubbers, silicone rubbers, plasticized trifluoro-monochloroethylene (Kel-F), etc., the exact material being chosen according to the requirements of the particular pumping problem. The flexures are best made of a high-carbon spring steel covered with a corrosion resistant paint or plastic, or of hard-tempered stainless steel. The seals may also be made of stainless steel or other metal, provided that they are formed with convolutions to give them flexibility in the axial direction.

The foregoing and other important aspects and features of my invention will be apparent from a consideration of the accompanying drawings, wherein:

FIG. 1 is a side view of a high-frequency, double-acting piston pump mounted on a base with an electrodynamic vibration motor as prime mover, with one of the flexure supports shown cut away for clarity;

FIG. 2 is a top view substantially along the line 22 of FIG. 1;

FIG. 3 is a sectional view substantially along the line 33 of FIG. 1;

FIG. 4 is a sectional view substantially along the line 44 of FIG. 1;

FIG. 5 is a sectional view substantially along the line 5-5 of FIG. 4;

FIG. 6 is a median sectional view substantially along the line 6--6 of FIG. 3, the flexure supports being shown cut away for clarity;

9-9 of FIG. 6; and

FIG. is a median sectional view of a bellows-type seal boot.

With reference to FIGS. 1 to 6, the pump P and prime mover M are mounted on a channel-shaped base or support 1, to which are bolted four upright posts 2 (FIGS. 1 and 3) and channel-section girders 3 (FlGS. 3 to 5) which are in the form of an inverted U. The upper ends of posts 2 are bolted by cap screws 4 to horizontal angle brackets 5, one bracket being located on each side of the pump unit. Each angle bracket 5 is secured by screws 6 10 its respective side of a cylinder block 7 (FIGS. 1 and 6).

Each of the girders 3 supports the upper end of one flexure assembly comprising two short, fiat leaf springs of high transverse stiffness 8 and 9 which are bolted in a horizontal position to the underside of channel 3 by cap screws 11 (FIG. 1), these springs 8 and 9 being separated by a solid metal spacer block 11 and extending in parallel overlying spaced relation. At their outer ends the springs 8 and 9 are bolted to a similar spacer block 12 (FIG. 1) by cap screws 13, which fit into tapped holes in nut plate 14. Secured to the outer vertical surface of spacer block 12 is the upper end of a main flexure (leaf spring) 15, the general plane of which extends at right angles to the general plane of the springs 8 and 9. The flexure 15 is clamped to the block 12 by a single cap screw 16 which passes through washer plate 17 and flexure 15 into a tapped hole in block 12. At the center of flexure 15 there is a hole through which passes the threaded shank of a special elongated plug 18 extending at right angles to the general plane of the spring 15. Between the hexagonal shoulder of plug 18 and a narrow hex nut 19 are two square washer plates 20 and 21 which confine flexure 15 and define the upper and lower horizontal lines along which bending of flexure 15 may occur. Nut 19 is tightened securely so as to clamp the parts in place during vibration. The lower end of each flexure 15 is clamped to a set of parts identical to those at its upper end, with the exception that the lower assembly of short flexures is secured directly to base 1 vertically below the place of attachment of the upper parts to channel 3.

Each flexure assembly is so designed that the stiffness against vertical displacements in the plane of the flexure 15 is very high, being in the order of 50,000 to 100,000 pounds per inch. The stiffness against horizontal displacements in the plane of flexure 15 is several orders of magnitude higher. Because of the two short flexures 8 and 9 used at each end, the blocks 12 can undergo no rotation, but only a vertical translation, so that when plug 18 undergoes a horizontal displacement of up to 4 inch from rest position, the concomitant shortening of the vertical length of flexure 15 (which amounts to no more than 0.020 inch in most cases) can occur without any change in the direction of flexure 15 at its ends. Furthermore, this shortening induces only minor tensile stresses in flexure 15 by comparison with the bending stresses caused by the transverse horizontal deflection at the center of the flexure 15. The advantage of this spring suspension is that it permits a true rectilinear motion of the members which are suspended between the centers of flexures 15, but without the severe changes of stiffness with increasing deflection which would occur if the upper and lower ends of flexures 15' were rigidly clamped.

The stiffness of the flexure system against horizontal motions on an axis passing through the centers of flexures 15 is readily and accurately computed from theory, and it is chosen so that the natural frequency of the moving system of the pump is equal or nearly to the frequency of operation. This removes the inertia-force load from the prime mover M which is itself designed so that its own natural frequency coincides with the operating frequency. The prime mover M is of the electrodynamic type having an armature shaft 25, such as shown in my copending application Serial No. 553,015, filed December 14, 1955, now Patent No. 2,936,713, to which reference may be had for a more complete disclosure.

Between the inner flanges of channels 3, at their upper portions, is fastened a stiffening rod 22 (FIGS. 1 and 2) resistance.

which adds rigidity to the flexure-supporting structure. Supported between plugs 18, with its outer ends threaded and screwed tightly into blind tapped holes in plugs 18, is a straight rod or shaft 30 (FIGS. 3, 6 and 7 to 9), having a circular cross section. Near each of its outer ends rod 30 passes through a long circular bushing 31 (FIGS. 3 and 6) which is coaxial with rod 30 and which has an inside diameter slightly larger than the outside diameter of rod 30, so that no contact occurs between rod and bushing.

The bushing 31 is press-fitted or otherwise secured coaxially within a larger sleeve 32, having three distinct diameters. The largest of these diameters, at the inner end of sleeve 32, fits freely into a mating hole at either end of cylinder block 7. The adjacent diameter, somewhat smaller, is either press-fitted or welded into a mating hole in a rectangular end plate 33, which is fastened to the end of cylinder block 7 by screws 34, which serve to compress a thin gasket 35 between plate 33 and block 7. The screws 34 are set into blind tapped holes in block 7, so that no leakage via the screw threads can occur. The smallest diameter of sleeve 32, at its outer end, is the same as the inside diameter of an elastomer sealboot 36, which is tightly clamped around sleeve 32 by a screw-and-band clamp 37. The other end of boot 36 is similarly clamped by another clamp 38 to a smooth circular portion at the inner end of plug 18. Because there are no openings to the outside, no leakage of the liquid being pumped can take place from the space within boot 36 to the outside environment.

During the reciprocating motion of shaft 30, the boots 36 are alternately elongated and compressed, the deformation, as a fraction of the normal length, being chosen so that the internal stresses in the elastomer are low, and the operating life correspondingly long. Because the space within each boot is essentially exposed to the. average pressure within the pump chambers 40 and '41, the boot is subjected to radial deformation outward on this account.

Secured to the. center of rod 30 by setscrews 42 is a cylindrical smooth walled piston 43 (FIGS. 6 to 9) having a circular cross section. Piston 43 is coaxial with shaft 30.and with the smooth walled cylinder bore in block 7, there being a small annular clearance space around piston 43 so that no contact occurs between piston and bore. The clearance area is chosen so that it is one per cent or less of the piston area, and the bore length made as great as practical, in order to minimize leakage flow past the piston. It will be noted that the cylinder bore, piston and enlarged pump chambers 40 and 41, which as shown in FIG. 6 are at the oposite ends of the piston, are constructed and arranged so that the leakage path between the chambers 40, 41 is' unconvoluted or in other words the leakage path between the chambers at the opposite ends of the piston is formed entirely by the radial spacing of the piston and bore, and the annular space between the piston and bore opens into and communicates directly with the chambers at the opposite ends of the piston. In this connection, it will be noted that the piston is provided with a rectangular longitudinal cross section in any plane passing through the piston axis.

As mentioned above, t each end of the cylinder bore, there is a somewhat enlarged liquid chamber '40 or 41 (FIG. 6), the chamber '40 communicating with an intake valve port 46 and a discharge valve port 44, while chamber 41 communicates with intake port 47 and discharge port 45. Each valve port is a circular opening in a sleevetype valve seat 50 to 53, preferably made of a'suitable plastic in order to suppress valve noise and insure wear These valve seats are lightly press-fitted into corresponding holes in the intake-valve block 49 and the discharge-valve block 48.

Each valve port is closed against downward liquid flow by means of a stack of flat leaf springs or reeds 54 to 57, each stack being composed of one long base reed with several reeds of progressively shorter length, (but the."

are blind. so as to prevent leakage to the outside.

or gas bubbles.

same thickness) above it, the whole assembly being clamped between pressure plates at its fixed end by screws set into tapped holes in. the valve block. The theory of such rectifying valves is fully described in my co-pending application Serial No. 636,597, filed January 28, I957, now abandoned. However, one refinement is added her, namely, the use of several reeds of different lengths. The advantage of this construction is that for a given stress level in all the reeds at maximum deflection the natural frequency of the composite valve is higher than for a single reed or a stack of several reeds all of the same length. The projection of each valve seat above the valve block is designed so that when closed each valve stack is under a slight preload, and this causes, the base reed to close the port against static backfiow, while insuring that all the reeds remain in contact with one another at all times.

The intake-valve block 49 is formed with internal passages having sloping surfaces, so that there is nowhere a horizontal surface or dome which can trap air Similarly, the discharge-valve block has no horizontal surfaces or domes in its internal passages. Interposed between valve block 48 and cylinder block 7 is a thin, flat gasket 71 which prevents leakage across the interface. Similarly, a gasket 72 is interposed between valve block 49 and cylinder block 7. Fastened below block 49 is an intake header 61, which serves to connect intake nipple 63 with valve ports 46 and 47. Nipple 63 is either press-fitted, screwed (via pipe threads), or welded into header 61. Between header 6i and block 49 is another thin, flat gasket 73 to prevent lea-kage across this interface. Connected securely to nipple 63 by means of a suitable clamp 76 is a flexible hose or tube 77 which carries the entering flow of liquid to the pump.

Directly above discharge-valve'block 48 is fastened a discharge header which serves to collect the pulsating flow which emerges alternately from valve ports 44 and 45, and to guide this flow to discharge nipple 62. Nipple 62,,which is securely fastened into header 60, has clamped to it by means .of clamp 74 a flexible discharge hose which carries the emergent flow to the external apparatus (not shown). Between header 6t) and valve block 48 is a gasket 70 to prevent leakage across the interface. Note that the interior space of header Gtlneed not be provided with a sloping roof, since the accumulation of -air or gas bubbles therein does no harm, and may in fact help to eliminate the pulsations from the liquid flow through hose 75.

As shown in FIGS. 7, 8 and 9, the header 6% and valve block 48 are secured to each other and to cylinder block 7 by long screws 80, which are regularly spaced near and around the periphery of members 7, 48 and 6t). These screws pass through clearance holes in parts 48,

60, 7.0 and 71 into tapped holes in cylinder block 7.

Note that wherever these tapped holes might emerge into any of the liquid chambers if carried through, the holes The central portion of block 48 is compressed against gasket 71 and block 7 by shorter screws 81 which pass through clearancev holes in parts 48. and 71 into tapped holes in block 7. If desired, these holes can run through into the cylinder bore, since the negligible leakage which might occur throughout them cannot escape to the outside environment. (Of course, the screws must not project into the bore.) In an exactly similar fashion, intake fheader 61 and valve block 49' are fastened to each other and to cylinderblock 7 by long screws 82, while screws Into a blind tapped hole in the outer end of the plug 18 which. is opposite prime mover M there is screwed the threaded end of a slender drive link 85 (FIGS. 1, 2 and 6) which carries a locknut 86 preventing this threaded shank from working loose when in motion. The other end. of link 85 is similarly fastened into the end of the armature shaft 25 of prime mover M by means of another threaded shank and locknut 87. The purpose of this drive link is to provide a mechanical connection between pump and driver that is sufiiciently rigid against axial compression and tension to transmit the requisite driving force, but sufiiciently flexible against transverse bending to permit small misalignments between the armature shaft of the driver and rod 39 of the pump, without inducing large bending stresses in either shaft.

In operation of the pump, the sequence of events is as follows: Assuming that piston 43 is instantaneously moving to the right in FIG. 6, the volume of chamber 40 is then being increased, and hence the pressure therein is decreasing. Valve 56 permits liquid to enter chamber 40 from chamber 66 via valve port 46, and simultaneously valve 54 prevents any liquid from entering 49 from chamber 65 via port 44. At the same time, the volume of chamber 41 is being reduced, and hence some of the liquid therein is forced out into chamber 65 via valve port 45. While this takes place, no liquid is allowed to escape from chamber 41 via port 47, because valve 57 closes this port. A small amount of liquid will leak from chamber 41 back to chamber 40 via the clearance space around piston 43 in the cylinder bore, but, as has been previously explained, this is an insignificant fraction of the total piston displacement per stroke. Also, a negligible amount of liquid may escape from 41 into the space within seal boot 36 on the same side as chamber 41.

On the return stroke (to the left), the events in chambers 4t and 41 will be interchanged, with liquid entering chamber 41 via port 47 and leaving chamber 40 via port 44. Thus, there is a pulsating flow of liquid from intake nipple 63 to discharge nipple 62, with two pulsations of the flow per cycle of piston displacement. To those who are familiar with electrical technology, it will be apparent that valves 54, 55, 56 and 57 are connected in the form of a single-phase, full-wave bridge, so as to rectify the alternating motion of the liquid which occurs in chambers 40 and 41.

The modification shown in FIG. is a variation of the seal boot wherein a bellows 95) is used in place of the cylindrical elastomer tube previously illustrated. If this bellows is made of an elastomer, then it may be clamped at its ends by clamps 91 and 92, but if the bellows is of metal, then the clamping means must consist of suitable end flanges and gaskets such as are customarily used with metal bellows. In either case, the volume of the space within the seal boot or bellows may be almost entirely filled with a stack of elastomer washers 93 mounted about the rod 30. The free length of this stack is chosen so that when the stack is installed it is compressed statically by an amount which exceeds the single amplitude of the motion of the shaft. Thus, the stack of washers is never entirely relieved of compression, and the washers always remain in contact with one another and with the pump members at each end of the stack. If the static compression of the stack is a suitably small fraction of the free length, then the compressive stresses in the stack during operation are kept low enough to insure long service life.

In summary, the advantages of my invention are as follows: first, the pump is constructed in such a way that no leakage of the liquid being pumped can occur between the interior of the pump and the outside environment; second, because theentire moving system is elastically suspended, without sliding contact between the moving and stationary parts, wear on the moving parts due to friction is eliminated; and third, because the external discharge is governed by the frequency and amplitude of the piston motion, the discharge may be controlled by electrical means acting on the prime mover, rather than by expensive flow-control valves.

I claim: 1 Y

1. In a liquid pump of the class described, a housing having a smooth walled bore, a smooth walled piston,

means including a piston rod to dispose said piston within said bore, said piston dividing the housing into two chambers located respectively at the opposite ends of the piston, the circumferential wall of said piston being radially inwardly spaced relative to the wall of said bore to provide a liquid leakage path past said piston and from one chamber to the other chamber, said leakage path formed by the radial spacing of the piston and bore opening at its opposite ends directly into said chambers, means for introducing liquid into each of said chambers including means for impeding liquid flow out of the chambers, means for conducting liquid out of said chambers including means for impeding back flow of liquid into said chambers, and means for reciprocating said piston longitudinally of said bore at a frequency of between 20 and 300 cycles per second.

2. In a liquid pump of the class described, a housing having a smooth walled bore, a smooth walled piston disposed within said bore and dividing the housing into a pair of chambers, the piston being cross sectionally smaller than said here to provide a substantial space between the circumferential walls of the piston and bore, a piston rod extending coaxially from the piston and supported in spaced out-of-contact relation to the housing, deformable sealing means between the rod and housing to prevent liquid flow externally of the housing around said piston rod, means for introducing liquid into each of said chambers including means for impeding liquid fiow out of the chambers, means for conducting liquid out of said chambers including means for impeding back flow of liquid into said chambers, and means for reciprocating said piston longitudinally of said bore at 21 frequency of between 20 and 300 cycles per second.

3. In a liquid pump of the class described, a housing having a smooth walled bore, a smooth walled piston disposed within said bore and dividing the housing into a pair of chambers, the piston being cross sectionally smaller than said bore to provide a substantial space between the circumferential walls of the piston and bore and provide a liquid leakage path between said chambers which is formed entirely by the radial spacing of said piston and bore and which opens directly into said chambers, a piston rod extending coaxially from the piston and supported in spaced out-of-contact relation to the housing, deformable sealing means between the rod and housing to prevent liquid flow externally of the housing around said piston rod, means for introducing liquid into each of said chambers including means for impeding liquid flow out of the chambers, means for conducting liquid out of said chambers including means for impeding back flow of liquid into said chambers, and means for reciprocating said piston longitudinally of said bore at a frequency of between 20 and 300 cycles per second.

4. In a liquid pump of the class described, a housing having a smooth walled bore, a smooth walled piston, means including a piston rod to dispose said piston within said bore, said piston dividing the housing into two chambers located respectively at the opposite ends of the piston, the circumferential wall of said piston being radially inwardly spaced relative to the wall of said bore to provide a liquid leakage path past said piston and from one chamher to the other chamber, the piston having a substantially rectangular longitudinal cross section in any plane taken through the piston axis to provide that the leakage path formed by the radial spacing of the piston and bore communicates directly with said chambers, means for introducing liquid into each of said chambers including means for impeding liquid flow out of the chambers, means for conducting liquid out of said chambers including means for impeding back flow of liquid into said chambers, and means for reciprocating said piston longitudinally of said bore at a frequency of between 20 and 300 cycles per second.

5. A liquid pump of the class described comprising a housing having a smooth walled bore, a piston rod extending coaxially of and within said bore and mounted for oscillatory movement longitudinally of the bore, said rod being spaced in out-of-contact relation to said housing, at least one end of said rod projecting outwardly of said housing, a smooth walled piston secured to said rod within said bore, said piston dividing said housing into two chambers, the circumferential wall of the piston being spaced radially inwardly of the wall of said bore to provide a liquid leakage path past said piston and between said chambers, said leakage path opening directly into said chambers, each of said chambers having an inlet port and an outlet port, inlet manifold means connecting the inlet ports, outlet manifold means connecting the outlet ports, rectifying valve means in said inlet ports operative to permit a liquid to flow only into the respective chambers, rectifying valve means in said outlet ports operative to permit a liquid to flow only from the respective chambers, a tubular member enveloping the projecting end of said rod and providing a fluid-tight seal between said projecting end of said rod and the housing, and means acting on said projecting end for oscillating said piston at a frequency of between 20 and 300 cycles per second.

6. A liquid pump of the class described comprising a housing having between its ends a cylindrical bore, sleeves at the opposite ends of said ho'using and coaxial with said bore, a piston rod extending in radially spaced coaxial relation relative to said bore and said sleeves, one end of said rod extending outwardly beyond the adjacent end of said housing, a piston secured to said rod within said bore, said piston dividing said housing into two chambers, the side wall of said piston being in out-of-contact relation to the wall of said bore, each of said chambers having an inlet port and an outlet port, means forming an inlet manifold secured to said housing and connecting the inlet ports, means forming an outlet manifold connecting the outlet ports, an inlet duct connected with the inlet manifold, an outlet duct connected with the outlet manifold, rectifying valve means in said inlet ports operative to permit a liquid to flow only into the respective chambers, rectifying valve means in said outlet ports operative to permit a liquid to flow only from the respective chambers, sealing'means providing a fluid-tight seal for said one end of the rod' including a tubular member having expandable and contractiole side walls telescopically surrounding the outwardly projecting one end of said rod, means connecting one end of said tubular member to the housing, means connecting the other end of the tubular member to said rod for movement therewith, and means acting on said one end of the rod for oscillating said piston at a frequency between 20 and 300 cycles per second. e V

7. .A liquid pump as set forthin claim 6 wherein the sealing means for the projecting end of said rod comprises a tubular elastomer clamped at one end to the outer end of one of said sleeves, and a metallic plug carried for movement by said one end of the rod and connected to the other end of said tubular elastomer to seal said other end of the elastomer and connect the same to said one' end of the rod for movement therewith.

8. A liquid pump as set forth in claim 6 wherein the sealing means for the projecting end of said rod comi nected to the other end of said bellows to close said other L end' of the bellows and connect the same for movement with said rod.

9. A liquid pump as set forth in claim 8, wherein a plurality of resilient elastomeric washers are circumposed about the projecting end of said rod within the confines of the bellows,.the Washers normally being under a compressive stress.

10. A liquid pump of the class described comprising a fixed housing, a support for the housing, a reciprocable rod extending through said housing, elastic means supporting the ends of said rod for substantially purely axial movement of said rod including a pair of vertically disposed fiat leaf springs each having a general plane extending at right angles to the axis of said rod, said springs being respectively clamped to the ends of said rod for movement therewith and being clamped to said support at points spaced from said rod, a piston supported by said rod intermediate its ends within said housing with its side wall portions in closely spaced out-of-contact relation to the adjacent inner walls of said housing, said piston dividing said housing into two chambers, each chamber having an inlet port and an outlet port, means forming an inlet manifold secured to said housing and connecting the inlet ports, means forming an outlet manifold secured to said housing and connecting the outlet ports, an inlet duct connected with the inlet manifold, an outlet duct connected with the outlet manifold, rectifying valve means in said inlet ports operative to permit a liquid flow only into the respective chambers, rectifying valve means in said outlet ports operative to permit a liquid flow only from the respective chambers, and means acting on one end of said rod for oscillating said piston.

11. A liquid pump as set forth in claim 10 wherein the opposite ends of said rod are secured to the centers of the vertically disposed leaf springs, a pluralityjof relatively short horizontally extending leaf springs each having a general plane extending at right angles to the general plane of the vertically disposed leaf springs, the upper and lower ends of each of said vertically disposed leaf springs being supported by said horizontally extending leaf springs, one end of each of said horizontally extending leaf springs being connected to one of said vertically disposed leaf springs and the other end of each of said horizontally extending leaf springs being fixed relative to the housing.

12. In a pump of the type having the housing provided with a horizontally extending bore, a piston reciprocably received in said bore in radially spaced relation to said bore, a piston rod extending from the piston and coaxially of said bore, and means to support the piston rod from movement coaxialy of said bore including a vertically arranged flat leaf spring extending at right angles to the axis of said rod adjacent one end thereof, means connecting the center of said leaf spring to said one end of the rod for movement therewith, and means supporting the leaf spring with its center aligned with the axis of said bore including a pair of relatively short horizontally arranged flat leaf springs disposed in parallel overlying spaced relation adjacent each of opposite ends of said vertically arranged leaf spring, means anchoring one end of the springs of each pair of said horizontally arranged springs in fixed relation to said housing, and means rigidly connecting the other ends of the springs of each pair of said horizontally arranged leaf springs to each other and to said vertically arranged leaf spring.

References Cited in the file of this patent UNITED STATES PATENTS 

1. IN A LIQUID PUMP OF THE CLASS DESCRIBED, A HOUSING HAVING A SMOOTH WALLED BORE, A SMOOTH WALLED PISTON, MEANS INCLUDING A PISTON ROD TO DISPOSE SAID PISTON WITH IN SAID BORE, SAID PISTON DIVIDING THE HOUSING INTO TWO CHAMBERS LOCATED RESPECTIVELY AT THE OPPOSITE ENDS OF THE PISTON, THE CIRCUMFERENTIAL WALL OF SAID PISTON BEING RADIALLY INWARDLY SPACED RELATIVE TO THE WALL OF SAID BORE TO PROVIDE A LIQUID LEAKAGE PATH PAST SAID PISTON AND FROM ONE CHAMBER TO THE OTHER CHAMBER, SAID LEAKAGE PATH FORMED BY THE RADIAL SPACING OF THE PISTON AND BORE OPENING AT ITS OPPOSITE ENDS DIRECTLY INTO SAID CHAMBERS, MEANS FOR INTRODUCING LIQUID INTO EACH OF SAID CHAMBERS INCLUDING MEANS FOR IMPEDING LIQUID FLOW OUT OF THE CHAMBERS, MEANS FOR CONDUCTING LIQUID OUT OF SAID CHAMBERS INCLUDING MEANS FOR IMPEDING BACK FLOW OF LIQUID INTO SAID CHAMBERS, AND MEANS FOR RECIPROCATING SAID PISTON LONGITUDINALLY OF SAID BORE AT A FREQUENCY OF BETWEEN 20 AND 300 CYCLES PER SECOND. 